Method of input power on/off discrimination for an automatic vehicular transmission system

ABSTRACT

A method of input power on/off discrimination for an automatic vehicular transmission system having an input shaft to which an output torque of an engine is transmitted, the method being adapted, for example, for the selection of the transmission control logic of the transmission system. The output torque of the engine transmitted to the input shaft is detected. If the detected output torque takes a value such that the rotating speed changing rate of the input shaft during transmission control is higher than a predetermined target value, it is concluded that the engine is in the power-on state. Preferably, the output torque of the engine is transmitted to the input shaft of the transmission system through a driving force transmission apparatus capable of detecting transmission torque. In this case, the speed changing rate of the engine and the transmission torque of the driving force transmission apparatus are detected, and the detected transmission torque and the product of the detected engine speed changing rate and a predetermined value are added so that the resulting sum is used as the value of the output torque of the engine. The power-on or -off state is discriminated by the output torque thus calculated.

BACKGROUND OF THE INVENTION

The present invention relates to a method of input power on/offdiscrimination for an automatic vehicular transmission system, and moreparticularly, to a method of input power on/off discrimination based onan output torque of an engine transmitted to an input shaft of anautomatic transmission system, in which the result of discrimination issuitably used for the selection of the transmission control logic of thetransmission system.

In a conventional transmission control method for anelectronically-controlled, automatic vehicular transmission system, thetransmission control logic varies depending on the power on/off state ofthe input torque of the transmission system. In an operation of shift-upfrom a lower gear ratio mode to a higher one, for example, if the inputtorque is in a power-on state, a frictional engagement element on therelease side is first gradually disengaged so that the rotating speed ofan input shaft of the transmission system is temporarily increased to alevel a little higher than that of the input shaft speed immediatelybefore the start of transmission control. Thereafter, a frictionalengagement element on the connection side is caused to start engagement,and its engaging force is then adjusted. The input shaft speed isgradually reduced toward a rotating speed for the establishment of thehigher gear ratio mode, while its changing rate is being adjusted to apredetermined value. When the input shaft speed attains the speed forthe higher gear ratio mode, the engagement of the connection-sideengagement element is completed.

If the input torque is in the power-off state at the time of theshift-up operation, on the other hand, the release-side frictionalengagement element is disengaged on delivery of a shift command, whilethe connection-side engagement element is kept stand-by at a positionjust short of an engagement start position until the rotating speed ofthe input shaft is lowered to a predetermined level. When the inputshaft speed is lowered to the predetermined level, the connection-sideengagement element is caused to start engagement, and its engaging forceis enhanced gradually. The input shaft speed is gradually reduced towardthe rotating speed for the establishment of the higher gear ratio mode,while its changing rate is being adjusted to the predetermined value.When the input shaft speed attains the speed for the higher gear ratiomode, the engagement of the connection-side engagement element iscompleted.

In an operation of shift-down from a higher gear ratio mode to a lowerone, moreover, if the input torque is in the power-on state, therelease-side frictional engagement element is first graduallydisengaged. The rotating speed of the input shaft is gradually increasedtoward a rotating speed for the establishment of the lower gear ratiomode, while its changing rate is being adjusted to a predeterminedvalue. Since the engine is in its power-on state, the input shaft speedincreases by itself if only part of the friction torque of therelease-side engagement element is removed. After the input shaft speedis temporarily increased to a level a little higher than that of thespeed for the lower gear ratio mode, the connection-side engagementelement is caused to start engagement, while keeping this speed level ofthe input shaft, and its engaging force is increased gradually. Thus,the engagement of the connection-side engagement element is completed.

If the input torque is in the power-off state at the time of theshift-down operation, on the other hand, the release-side frictionalengagement element is gradually disengaged so that the rotating speed ofthe input shaft is temporarily lowered to a level a little lower thanthat of the input shaft speed immediately before the start oftransmission control. Thereafter, the connection-side frictionalengagement element is caused to start engagement, and its engaging forceis increased gradually. The input shaft speed is gradually increasedtoward the rotating speed for the establishment of the lower gear ratiomode, while its changing rate is being adjusted to the predeterminedvalue. When the input shaft speed attains a speed level a little lowerthan the speed for the lower gear ratio mode, the engagement of theconnection-side engagement element is completed.

In the aforementioned transmission control method using the transmissioncontrol logic varying according to the power on/off state of the inputtorque, it is essential to discriminate the power on/off statecorrectly. Conventionally, the discrimination of the power on/off stateis based on the polarity, i.e., acceleration or deceleration of theengine output. Such a conventional method is subject to the followingdrawbacks.

In the discrimination method described above, if the operation is in anup-shift mode, for example, the power-off state is detected even whenthe engine output is only slightly decelerating or negative. Therefore,the release-side frictional engagement element is disengaged at once,while the connection-side engagement element is kept at the stand-byposition until the rotating speed of the input shaft is lowered to thepredetermined level. Thus, the input shaft speed cannot readily lower,so that the start of the engagement of the connection-side frictionalengagement element is delayed.

In the case of a down-shift mode, moreover, the power-on state isdetected even when the engine output is only slightly positive.Accordingly, automatic increase of the rotating speed of the input shaftis awaited, and the input shaft speed cannot readily increase. Thus, thecompletion of the engagement of the connection-side engagement elementis delayed.

OBJECT AND SUMMARY OF THE INVENTION

The primary object of the present invention is to provide a method ofinput power on/off discrimination for an automatic vehiculartransmission system, capable of proper and secure discrimination of thepower on/off state of the input torque of a transmission system, and ofquickly finishing transmission control by selecting the transmissioncontrol logic of the transmission system in accordance with the resultof the discrimination.

According to the present invention, there is provided a method of inputpower on/off discrimination for an automatic vehicular transmissionsystem having an input shaft to which an output torque of an engine istransmitted, the method being adapted, for example, for the selection ofthe transmission control logic of the transmission system. The outputtorque of the engine transmitted to the input shaft is detected. If thedetected output torque takes a value such that the rotating speedchanging rate of the input shaft during transmission control is higherthan a predetermined target value, it is concluded that the engine is ina power-on state. If the detected output torque takes a value such thatthe rotating speed changing rate is lower than the predetermined targetvalue, on the other hand, it is concluded that the engine is in apower-off state.

Preferably, the output torque of the engine is transmitted to the inputshaft of the transmission system through a driving force transmissionapparatus capable of detecting transmission torque. In this case, thespeed changing rate of the engine and the transmission torque of thedriving force transmission apparatus are detected, and the detectedtransmission torque and the product of the detected engine speedchanging rate and a predetermined value are added so that the resultingsum is used as the value of the output torque of the engine. Thepower-on or -off state is discriminated by the output torque thuscalculated.

The present invention is based on the understanding that the poweron/off discrimination for the input torque should be effected dependingon whether the rotating speed of the input shaft of the transmissionsystem can be changed at a predetermined target changing rate when allthe frictional engagement elements concerned in the transmission controlare disengaged. Thus, even though the engine output is positive, forexample, the power-on state cannot be detected unless the input torquetakes a value such that the rotating speed changing rate of the inputshaft of the transmission system, at the time of transmission control,is higher than the predetermined target value. In consequence, thetransmission control does not require any long time.

The above and other objects, features, and advantages of the inventionwill be more apparent from the ensuing detailed description taken inconneciton with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a block diagram showing an outline of an automatictransmission with a torque converter, to which the present invention isapplied;

FIG. 2 is a gear train diagram showning part of the internal arrangmentof a gear transmission 30 shown in FIG. 1;

FIG. 3 is a hydraulic circuit diagram showing part of the internalarrangement of a hydraulic circuit 40 shown in FIG. 1;

FIG. 4 is a flow chart for a main routine illustrating hydraulic controlprocesses executed during transmission control by means of atransmission control unit (TCU) 16 shown in FIG. 1;

FIG. 5 is a timing chart illustrating the way pulse signals from anengine speed (Ne) sensor 14, used to calculate an engine speed Ne, areproduced;

FIG. 6 is a shift map showing transmission control regions defined by athrottle valve opening and a transfer drive gear speed No;

FIG. 7 is a flow chart of a power on/off decision routine executed bymeans of the transmission control unit (TCU) 16;

FIGS. 8 to 12 are flow charts illustrating processes of hydrauliccontrol executed in a power-on up-shift mode by means of thetransmission control unit (TCU) 16;

FIGS. 13(a), 13(b) and 13(c) are timing charts showing time-basedtransitions of a turbine speed Nt and the transfer drive gear speed Noand transitions of the respective duty factors of release- andconnection-side solenoid valves, used in the power-on up-shift mode;

FIGS. 14, 15 and 16 are flow charts illustrating processes of hydrauliccontrol executed in a power-on down-shift mode by means of thetransmission control unit (TCU) 16;

FIGS. 17(a), 17(b) and 17(c) are timing charts showing time-basedtransitions of the turbine speed Nt and the transfer drive gear speed Noand transitions of the respective duty factors of the release- andconnection-side solenoid valves, used in the power-on down-shift mode;

FIGS. 18, 19 and 20 are flow charts illustrating processes of hydraulicontrol executed in a power-off up-shift mode by means of thetransmission control unit (TCU) 16;

FIGS, 21(a), 21(b) and 21(c) are timing charts showing time-basedtransitions of the turbine speed Nt and the transfer drive gear speed Noand transitions of the respective duty factors of the release- andconnection-side solenoid valves, used in the power-off up-shift mode;

FIGS. 22, 23 and 24 are flow charts illustrating processes of hydrauliccontrol executed in a power-off down-shift mode by means of thetransmission control unit (TCU) 16;

FIGS. 25(a), 25(b) and 25(c) are timing charts showing time-basedtransitions of the turbine speed Nt and the transfer drive gear speed Noand transitions of the respective duty factors of the release- andconnection-side solenoid valves, used in the power-off down-shift mode;and

FIG. 26 is a timing chart for illustrating time-based transitions of thethrottle valve opening, turbine shaft torque, and output shaft torqueused in a lift-foot up-shift mode.

DETAILED DESCRIPTION

FIG. 1 shows an outline of an electronically-controlled automaticvehicular transmission with a torque converter according to the presentinvention. In FIG. 1, an internal combustion engine 10, which is e.g. a6-cylindered engine, has a crankshaft 10a fitted with a flywheel 11. Oneend of a drive shaft 21 of a torque converter 20, for use as a drivingforce transmission device, is coupled mechanically to the crankshaft 10avia the flywheel 11. The torque converter 20 includes a casing 20a, apump 23, a stator 24, and a turbine 25. The pump 23 is coupled to theother end of the drive shaft 21 through an input casing 22 of theconverter 20, and the stator 24 is coupled to the casing 20a by means ofa one-way clutch 24a. The turbine 25 is connected to an input shaft 30aof a gear transmission 30.

In the present embodiment, the torque converter 20 is provided with adirect-coupled clutch of a slip-type, e.g., a damper clutch 28, which isdisposed between the input casing 22 and the turbine 25. Even when it isengaged or directly coupled, the damper clutch 28 allows a suitable slipbetween the pump 23 and the turbine 25 of the torque converter 20 whichare coupled mechanically and directly to each other. The slippage of thedamper clutch 28, i.e., the torque transmitted through the clutch 28, isexternally controlled by means of a hydraulic damper clutch controlcircuit 50.

The hydraulic damper clutch control circuit 50 includes a damper clutchcontrol valve 52 and a damper clutch control solenoid valve 54. Thesolenoid valve 54 is a normally-closed on-off valve, whose solenoid 54ais connected electrically to a transmission control unit (hereinafterreferred to simply as TCU) 16. The damper clutch control valve 52 servesto change an oil passage for operating oil to be supplied to the damperclutch 28, and to control the oil pressure acting on the clutch 28. Toattain this, the control valve 52 is composed of a spool 52a and aspring 52c. The spring 52c is contained in a left-end chamber 52b,fronted by the left end face, as illustrated, of the spool 52a, andserves to urge the spool 52a to the right of FIG. 1. The left-endchamber 52b is connected with a pilot oil passage 55 which communicateswith a pilot hydraulic source (not shown). The pilot oil passage 55 isconnected with a branch passage 55a which extends to the drain side. Thesolenoid valve 54 is located in the middle of the branch passage 55a.The level of a pilot oil pressure supplied to the left-end chamber 52bis controlled as the solenoid valve 54 is opened or closed. The pilotoil pressure from the pilot hydraulic source is also supplied to aright-end chamber 52d which is fronted by the right end face of thespool 52a.

When the pilot oil pressure acts on the left-end chamber 52b to causethe spool 52a of the damper clutch control valve 52 to move to the rightlimit position, a torque converter (T/C) lubricating oil pressure,supplied to the torque converter 20, is fed through an oil passage 56,the control valve 52, and the oil passage 57 into a hydraulic chamber,which is defined between the input casing 22 and the damper clutch 28.Thereupon, the damper clutch 28 is disengaged. On the other hand, whenthe pilot oil pressure is not fed into the left-end chamber 52b so thatthe spool 52a moves to the left limit position as illustrated, a linepressure from a hydraulic pump (not shown) is fed through an oil passage58, the control valve 52, and an oil passage 59 into an oil chamber,which is defined between the damper clutch 28 and the turbine 25. Thus,the damper clutch 28 is frictionally engaged with the input casing 22.When the duty factor (the ratio of the solenoid valve to the duty cycletime) Dc of the damper clutch solenoid valve 54 is controlled by meansof the TCU 16, the spool 52a is moved to the position where theresultant force of the urging force of the spring 52c and the pilot oilpressure acting on the left-end chamber 52b balances with the urgingforce of the pilot oil pressure acting on the right-end chamber 52d. Anoil pressure corresponding to the moved position is supplied to thedamper clutch 28, whereby a transmission torque Tc of the clutch 28 isadjusted to a required value.

The gear transmission 30 includes a gear train adapted for four forwardgear ratios and one backward gear ratio, for example. FIG. 2 is adiagram showing part of the arrangement of the gear transmission 30.First and second driving gears 31 and 32 are lossely fitted on the inputshaft 30a for rotation, and hydraulic clutches 33 and 34, for use asfrictional engagement elements for transmission control, are fixed tothat portion of the input shaft 30a between the driving gears 31 and 32.The driving gears 31 and 32 are adapted to rotate in one with the inputshaft 30a when they engage the clutches 33 and 34, respectively. Anintermediate transmission shaft 35, which extends parallel to the inputshaft 30a, is connected to a driving axle by means of a final reductiongear system (not shown). The intermediate transmission shaft 35 isfixedly fitted with first and second driven gears 36 and 37, which arein mesh with the first and second driving gears 31 and 32, respectively.When the clutch 33 engages the first driving gear 31, the rotation ofthe input shaft 30a is transmitted to the clutch 33, first driving gear31, first driven gear 36, and intermediate transmission gear 35. Thus, afirst transmission control mode (e.g., first gear ratio mode) isestablished. When the clutch 34 engages the second driving gear 32 afterthe clutch 33 is disengaged, the rotation of the input shaft 30a istransmitted to the clutch 34, second driving gear 32, second driven gear37, and intermediate transmission shaft 35. Thus, a second transmissincontrol mode (e.g., second gear ratio mode) is established.

FIG. 3 shows a hydraulic circuit 40 which supplies oil pressure to thehydraulic clutches 33 and 34. The hydraulic circuit 40 includes firstand second hydraulic control valves 44 and 46 and solenoid valves 47 and48. Spools 45 and 49 are slidably fitted in bores 44a and 46a of thefirst and second hydraulic control vavles 44 and 46, respectively. Thus,right-end chambers 44g and 46g are defined which are fronted by therespective right end faces of the spools 45 and 49. Springs 44b and 46b,which are contained in the chambers 44g and 46g, respectively, urgetheir corresponding spools 45 and 49 to the right of FIG. 3. The firstand second hydraulic control valves 44 and 46 are formed, respectively,with left-end chambers 44h and 46h which are fronted by the respectiveleft end faces of the spools 45 and 49. These chambers 44h and 46hcommunicate with the drain side by way of orifices 44i and 46i,respectively.

The solenoid valve 47 is a normally-open three-way valve which has threeports 47c, 47d and 47e. The valve 47 is composed of a vavle plug 47a, aspring 47b, and a solenoid 47f. The spring 47b serves to urge the valveplug 47a toward the port 47e, thereby closing the port 47e. When thesolenoid 47f is excited, it causes the valve plug 47a to move toward theport 47c against the urging force of the spring 47b, thereby closing theport 47c. The solenoid valve 48, on the other hand, is a normally-closedthree-way valve which has three ports 48c, 48d and 48e. The valve 48 iscomposed of a valve plug 48a, a spring 48b, and a solenoid 48f. Thespring 48b serves to urge the valve plug 48a toward the port 48c,thereby closing the port 48c. When the solenoid 48f is excited, itcauses the valve plug 48a to move toward the port 48e against the urgingforce of the spring 48b, thereby closing the port 48e. The respectivesolenoids 47f and 48fof the solenoid valves 47 and 48 are connected tothe output side of the TCU 16.

An oil passage 41, which extends from the aforesaid hydraulic pump (notshown), is connected to ports 44c and 46c of the first and secondhydraulic control valves 44 and 46. One end of oil passage 41a isconnected to a port 44d of the first hydraulic control valve 44, and thehydraulic clutch 33 is connected to the other end of the oil passage41a. One end of an oil passage 41b is connected to a port 46d of thesecond hydraulic control valve 46, and the hydraulic clutch 34 isconnected to the other end of the oil passage 41b. An oil passage 42,which extends from the aforesaid pilot hydraulic source (not shown), isconnected to ports 44e and 46e, communicating with the left-end chambers44h and 46h of the first and second hydraulic control valves 44 and 46,respectively, and also to the ports 47c and 48c of the solenoid valves47 and 48, respectively. The ports 47d and 48d of the solenoid valves 47and 48 are connected, respectively, to ports 44f and 46f whichcommunicate with the right-end chambers 44g and 46g of the first andsecond hydraulic control valves 44 and 46 by means of pilot oil passages42a and 42b, respectively. The ports 47e and 48e of the solenoid valves47 and 48 communicate with the drain side.

The oil passage 41 is used to supply the first and second hydrauliccontrol valves 44 and 46 with an operating oil pressure or linepressure, which is adjusted to a predetermined level by means of apressure-regulating valve (not shown) or the like. The pilot oil passage42 is used to supply the first and second hydraulic control valves 44and 46 and the solenoid valves 47 and 48 with a pilot oil pressure,which is adjusted to a predetermined level by means of anotherpressure-regulating valve (not shown) or the like.

When the spool 45 of the first hydraulic control valve 44 moves to theleft of FIG. 3, a land 45a of the spool 45, having so far been closingthe port 44c, allows the port 44c to open, so that the operating oilpressure is fed to the clutch 33 through the oil passage 41, the ports44c and 44d, and the oil passage 41a. When the spool 45 moves to theright, the port 44c is closed by the land 45a, while the port 44dcommunicates with a drain port 44j, so that the oil pressure inside theclutch 33 is discharged to the drain side. When the spool 49 of thesecond hydraulic control valve 46 moves to the left of FIG. 3, a land49a of the spool 49, having so far been closing the port 46c, allows theport 46c to open, so that the operating oil pressure is fed to theclutch 34 through the oil passage 41, the ports 46c and 46d, and the oilpassage 41b. When the spool 49 moves to the right, the port 46c isclosed by the land 49a, while the port 46d communicates with a drainport 46j, so that the oil pressure inside the clutch 34 is discharged tothe drain side.

Returning to FIG. 1, a ring gear 11a is fitted on the outer periphery ofthe flywheel 11, and is in mesh with a pinion 12a of a starter 12. Thering gear 11a has a predetermined number of the teeth (e.g., 110 teeth),and an electromagnetic pickup (hereinafter referred to as Ne sensor) 14is opposed to the ring gear 11a. The Ne sensor 14, which is used todetect the rotational speed Ne of the engine 10, as described in detaillater, is connected electrically to the input side of the TCU 16.

A turbine speed sensor (Nt sensor) 15, a transfer drive gear speedsensor (No sensor) 17, a throttle valve opening sensor (θt sensor) 18,an oil temperature sensor 19, etc. are also connected to the input sideof the TCU 16. The Nt sensor 15 and the No sensor 17 are used to detectthe respective speeds Nt and No of the turbine 25 of the torqueconverter 20 and a transfer drive gear (not shown). The θt sensor 18serves to detect the opening θt of throttle valve (not shown) which isdisposed in the middle of suction passage (not shown) of the engine 10.The oilk temperature sensor 19 is used to detect the temperature Toil ofthe operating oil discharged from a hydraulic pump (not shown).Detection signals from these sensors are supplied to the TCU 16.

The operation of the gear transmission with the aforementionedconstruction will now be described.

The TCU 16 contains therein memories, such as a ROM, RAM, etc., centralprocessing unit, I/O interface, counter, and the like. The TCU 16performs hydraulic control for the transmission control in the followingmanner, in accordance with a program stored in the memories.

The TCU 16 repeatedly executes a main program routine shown in FIG. 4with a predetermined cycle, e.g., 35-Hz cycle. In this main programroutine, various initial values, as mentioned later, are first read orset in step S10. Then, the TCU 16 reads the detection signals from thevarious sensors, including the Ne sensor 14, Nt sensor 15, No sensor 17,θt sensor 18, and oil temperature sensor 19, and stores the signalstherein (step S11). Thereafter, the TCU 16 calculates and storesnecessary parameter values for the transmission control, based on thedetection signals, in the following manner.

First, the TCU 16 calculates the engine speed Ne and its changing rateωe on the basis of the detection signal from the Ne sensor 14 (stepS12). The Ne sensor 14 delivers one pulse signal to the TCU 16 everytime it detects four of the teeth of the ring gear 11a during onerevolution of the gear 11a. Then, TCU 16 measures the period of time tp(sec) required for the detection of the last 9 pulses out of those pulsesignals which are supplied during one duty cycle, i.e., in 28.6 msec (35Hz), as shown in FIG. 5. Thereafter, the TCU 16 calculates the enginespeed Ne (rpm) in accordance with the following equation (1), and storesin the memories the obtained value as an engine speed (Ne)_(n) for thepresent duty cycle. ##EQU1##

Based on an engine speed (Ne)_(n-1) stored in the last duty cycle andthe engine speed (Ne)_(n) stored in the present duty cycle, the enginespeed changing rate ωe (rad/sec²) is calculated as follows and thenstored: ##EQU2## where there are relations ΔNe=(Ne)_(n) -(Ne)_(n-1) andT=(T1+T2)/2, and T1 and T2 are periods of time (sec) between therespective count end points of the periods tp in the last and presentduty cycles and between the respective count start points thereof,respectively.

Calculation of Turbine Shaft Torque Tt

Then, the TCU 16 proceeds to step S13, and calculates the net torque Teof the engine and the torque (hereinafter referred to as turbine shafttorque) Tt (kg·m) of the output shaft of the torque converter.

The relationships fo the friction torque Tb of the clutch on the releaseor connection side, obtained during the transmission control, to theturbine shaft torque Tt and the turbine speed changing rate ωt duringthe transmission control may be given as follows:

    Tb=a×Tt+b×ωt,                            (A1)

where a and b are constants which depend on the shift pattern or shiftschedule (type of transmission control), such as shift-up from firstgear ratio to second or shift-down from fourth gear ratio mode to third,the moments of inertia of various rotating parts, etc. As seen fromequation (A1), the clutch friction torque Tb, that is, the operating oilpressure for the clutches 33 and 34, can be set without receiving theinfluences of lowering of the engine performance, change of engine watertemperature, etc., if it is determined on the basis of the turbine shafttorque Tt and the turbine speed changing rate ωt. Empirical formulas anddata obtained in consideration of these circumstances may be readilyapplied to engines of different types.

If the turbine speed changing ratre ωt is expected to be subjected tofeedback control for a target value, despite the change of the turbineshaft torque Tt, what is required is not to correct the deviation of thechanging rate ωt from the target value afterward, but to increase ordecrease the friction torque Tb, i.e., the operating oil pressure of theclutches 33 and 34, by a margin corresponding to the variation of theturbine shaft torque Tt. By doing this, stable transmission control canbe effected with high follow-up performance, without requiring anysubstantial correction gain for feedback control.

If the time-based change of the turbine shaft torque Tt at the start ofthe generation of the friction torque of the connection-side clutch canbe estimated at the start of the transmission control, the clutchfriction torque can be changed while controlling the turbine speedchanging rate ωt around the target value, in accordance with equation(A1). Therefore, such a change of the torque Tt is empirically obtainedin advance. Based on empirical data thus obtained, the change of theturbine shaft torque Tt, at the start of the actual generation of thetorque of the connection-side clutch, is estimated. By applying theestimated value to equation (A1), the oil pressure supplied to theclutch can be adjusted to change the friction torque Tb so that thetarget value of the turbine speed changing rate ωt can be obtainedaccording to equation (A1). By doing this, the turbine speed changingrate ωt can be accurately controlled around the target value from thestart of the generation of the friction torque of the connection-sideclutch. Thus, the feeling of operation for the transmission control canbe improved.

Thereupon, the turbine shaft torque Tt is calculated according toequation (4), using the net engine torque Te calculated according toequation (3), and these calculated values are stored in the memories:##EQU3## Here Te is a net torque which is obtained by subtracting thefriction loss, oil pump driving torque, etc. from an average torqueproduced by the explosion of the engine 10, and C is a torque capacitycoefficient, which is read from a torque converter characteristic table,previously stored in the memories, in accordance with the ratio e(=Nt/Ne) of the turbine speed Nt to the engine speed Ne. After the speedratio e is first calculated according to the turbine speed Nt, detectedby means of the Nt sensor 14, and the engine speed Ne detected in theaforesaid manner, the torque capacity coefficient C is read from thememories in accordance with the calculated speed ratio e. I_(E) is themoment of inertia of the engine 10, which is fixed value set for eachengine type, and t is a torque ratio, which is also read from theaforesaid torque converter characteristic table in accordance with thespeed ratio e (=Nt/Ne).

Tc is the transmission torque of the damper clutch 28, which, in thedirect-coupled clutch of this slip type, is given by ##EQU4## where Pc,A, r, and μ are the supplied oil pressure, piston pressure receivingarea, friction area, and coefficient of friction, respectively, of thedamper clutch 38. Equation (5) can be obtained because the supplied oilpressure Pc of the damper clutch 28 is proportional to the duty factorDc of the damper clutch solenoid valve 54. In equation (5), a1 and b1are constants which are set according to the shift mode. The value Tccalculated by equation (5) is effective only if it is positive. If it isnegative, Tc is regarded as Tc=0.

The instantaneous values fo the net engine torque Te and the turbineshaft torque Tt, calculated and stored in this manner, can be calculatedon the basis of the engine speed Ne detected by the Ne sensor 14, theturbine speed Nt detected by the Nt sensor 15, and the duty factor Dc ofthe damper clutch solenoid valve 54. As seen from equations (3) and (4),moreover, the engine output torque Te is calculated including the term(I_(E) ×ωe), so that it hardly receives the influence of the turbinespeed changing rate ωt or friction torque Tb. Therefore, if the frictiontorque Tb is adjusted, that is, if the supplied pressure of the clutchis adjusted, in order to set the changing rate ωt to the target value,the turbine shaft torque Tt never changes. Thus, these two torquescannot interfere with each other, and therefore, never entail anyuncontrollable situations. In the middle of the transmission control, inparticular, if the friction torque Tb is adjusted so as to correctvariation of the turbine shaft torque Tt, which is caused by disturbanceattributable to acceleration work or the like, the aforesaidinterference cannot be entailed. Thus, the transmission control can beperformed with satisfactory responsiveness.

Then, in step S14, the TCU 16 determines the transmission gear ratiomode to be established in the gear transmission 30, on the basis of theopening θt of the throttle valve and the transfer drive gear speed No.FIG. 6 shows transmission control regions for the first transmissioncontrol mode (hereinafter referred to as first gear ratio mode) and thesecond transmission control mode (hereinafter referred to as second gearratio mode) one grade higher than the first mode. In FIG. 6, the fullline represents a boundary line between the regions for the first andsecond gear ratio modes, provided for the shift-up operation from thefirst gear ratio mode to the second. The broken line represents aboundary line between the first and second gear ratio mode regionsprovided for the shift-down operation from the second gear ratio mode tothe first. The TCU 16 determines the transmission gear ratio mode to beestablished according to the shift map of FIG. 6, and stores datarepresenting the predetermined mode in the memories in advance.

Power On/Off Discrimination

Then, the TCU 16 proceeds to step S15, and executes a power on/offdiscrimination routine. FIG. 7 is a flow chart showing this routine.First, a discrimination value Tto is set in step S151. The value Tto iscalculated as follows:

    Tto=a2×ωto=2π×a2×Ni,            (6)

where a2 and Ni are predetermined values previously set in accordancewith the shift schedule. The values a2 and Ni are negative in an upshift mode and positive in a down-shift mode. Then, the TCU 16determines whether the turbine shaft torque Tt, calculated in step S13,is greater than the discrimination value Tto (step S152). If theconclusion is YES, power-on shift is identified (step S153). If NO,power-off shift is identified (step S154). The TCU 16 stores thememories with the result of power on/off discrimination (step S155), andthen returns to the main routine shown in FIG. 4.

The above method of power on/off discrimination is based on thefollowing principle. Generally, equation (6) is obtained if the turbineshaft torque T, the turbine speed changing rate ωt for transmissioncontrol, and the clutch friction torque Tb are replaced by 0, ωto, andTto, respectively, in equation (A1) which represents the relationshipsof the value Tb to the values Tt and ωt. When no other elements exceptthe clutches are operating, the power on/off discrimination is effecteddepending on whether or not the turbine shaft torque Tt produced isgreat enough to ensure the target value ωto. Thus, the followingdrawbacks of the conventional discrimination method, in which the poweron/off discrimination depends simply on the polarity or accelerationsense of the engine output, can be eliminated.

More specifically, the transmission control system using differentlogics for power-on and -off states is subject to the followingdrawbacks.

(1) If the engine output is somewhat negative in the up shift mode, thepower-off state will be detected wrongly. As a result, theconnection-side friction element (clutch) is left disengaged, so thatthe transmission control cannot be completed.

(2) If the engine output is somewhat positive in the down-shift mode, onthe other hand, the power-on state will be detected wrongly.Accordingly, automatic increase of the rotating speed of the input shaftof the transmission is awaited, so that the connection-side frictionelement (clutch) fails to be connected. Thus, the transmission controlcannot be completed.

A lift-foot up-shift operation or a down-shift operation with a driver'sfoot on an accelerator pedal requires as fast power on/offdiscrimination as possible. The turbine shaft torque Tt used in theaforesaid power on/off discrimination is, so to speak, an imaginaryturbine shaft torque which is obtained by multiplying the net enginetorque Te, obtained from equation (3), by the torque ratio t. Thus, thepower on/off discrimination can be achieved more quickly than thediscrimination using an actual turbine shaft torque Tt' (=T×CNe² +Tc)which is obtained by omitting the term (I_(E) ×ωe) from equation (4).Thus, during the lift-foot up- shift operation, for example, the shockof speed reduction can be avoided, in a low gear ratio mode, if thereduction of the engine output is sensed as soon as possible so that therelease-side element (clutch) is disengaged without delay. Referring toFIG. 26, when the accelerator pedal is released so that the up shiftmode is established (FIG. 26(a)), the actual turbine shaft torque Tt'and the imaginary turbine shaft torque Tt change along the broken andfull lines of FIG. 26(b), respectively. Accordingly, the power-off statecan be detected at time t1 of FIG. 26(b) when the imaginary turbineshaft torque Tt is used, and at time t2 when the actual turbine shafttorque Tt' is used. Thus, with use of the imaginary torque Tt, thepower-off discrimination can be performed faster by Δt (=t2-t1) thanwhen the actual torque Tt' is used. Correspondingly, the release-sideelement can be disengaged more quickly, so that the shock of speedreduction can be avoided without a drop (see hatched region of FIG.26(c)) of the output shaft torque.

Returning to FIG. 4, the TCU 16 then determines whether the transmissioncontrol region to be established, which is determined in step S14, isdifferent from the result obtained in the preceding duty cycle. If thereis no difference, the program returns to step S11, and step S11 and thesubsequent steps are repeated again. If the transmission control regionis changed, on the other hand, a shift signal correspondikng to theshift pattern discriminated in steps S14 and S15 is outputted (stepS17), whereupon the program returns to step S11.

Hydraulic Control for Power-On Up-Shift

FIGS. 8 to 12 are flow charts showing processes of hydraulictransmission control for a power-on up-shift mode. Referring to FIG. 13,the processes of hydraulic control will be described in connection withthe case of shift-up operation from the first gear ratio mode to thesecond, by way of example.

In response to a shift signal for a power-on up-shift from the firstgear ratio mode to the second, the TCU 16 first calculates therespective initial duty factors D_(U1) and D_(U2) of the solenoid valves47 and 48 according to the following equations (8) and (9) (step S20).

    D.sub.U1 =a4×|Tt|+c4,              (8)

    D.sub.U2 =a5×|Tt|+c5,              (9)

where Tt is the turbine shaft torque value calculated and stored in stepS13 of FIG. 4 for each duty cycle, and a4, c4, a5, and c5 are constantsapplied to the case of the shift-up operation from the first gear ratiomode to the second.

Then, the TCU 16 adjusts the duty factory D_(LR) of the normally-opensolenoid valve 47 to the initial duty factory D_(U1) set in step S20,and delivers an output signal such that the solenoid valve 47 isoperated with the duty factor D_(LR). Thereupon, the first-speed clutch33, as a release-side frictional engagement element, starts to besupplied with an initial oil pressure corresponding to the initial dutyfactor D_(U1), so that a piston (not shown) of the clutch 33 isretreated toward a position just short of the position where the clutch33 slips (step S21; time t1 of FIG. 13(b)). Meanwhile, the TCU 16 setsthe duty factor D24 of the normally-closed solenoid valve 48 to 100%,and delivers an output signal such that the solenoid valve 48 isoperated with the duty factor D24. Thereupon, a piston of thesecond-speed clutch 34, as a connection-side frictional engagementelement, is advanced to a position just short of the position where theclutch 34 starts to engage (at time t1 of FIG. 13(c)), and an initialpressure supply period T_(S1) is set in a timer (step S22). This timermay be a hard timer contained in the TCU 16, or a so-called soft timerwhich times the supply period T_(S1) as the program is executed. Theinitial pressure supply period T_(S1) takes a predetermined value suchthat the piston of the connection-side clutch 34 can be advanced to thepredetermined position just short of the engagement start position ifthe clutch 34 is supplied with the operating oil pressure with 100-%duty factor all during the period T_(S1).

The TCU 16 waits until a predetermined period t_(D), i.e., one dutycycle (28.6 msec in this embodiment), terminates (step S23), and thenadds a predetermined duty factor ΔD1 to the duty factor D_(LR) set inthe preceding duty cycle, thus providing a new duty factor D_(LR). Then,the TCU 16 delivers an output signal such that the solenoid valve 47 isoperated with the duty cycle D_(LR) (step S24). The added predeterminedduty factor ΔD1 is set to a value such that the duty factor D_(LR) ofthe solenoid valve 47 increases at a predetermined rate (e.g., 4% persecond) (see the transition of the duty factor D_(LR) from time t1 to t2in FIG. 13(b)). The TCU 16 determines whether the initial pressuresupply period T_(S1), set in step S22, is over (step S25). If the periodT_(S1) is not over, the program returns to step S23, and steps S23, S24,and S25 are repeated.

If the conclusion of step S25 is YES, that is, when the second-speedclutch 34 advances to the predetermined position just short of theengagement position after the termination of the initial pressure supplyperiod T_(S1), the program proceeds to step S27 of FIG. 9. In step S27,the TCU 16 adjusts the duty factor D24 of the solenoid value 48 to apredetermined value D24min, and then delivers a driving signal such thatthe valve 48 is operated with the duty factor D24 (at time t2 of FIG.13(c)). The predetermined value D24min is a duty factor value such thatthe operating oil pressure supplied to the second-speed clutch 34 viathe second hydraulic control valve 46 is kept at a hold pressure withoutincreasing or decreasing. When the predetermined period t_(D) for oneduty cycle terminates (step S28), the TCU 16 adds the predetermined dutyfactor ΔD1 to the duty factor D_(LR) of the solenoid valve 47 set in thepreceding duty cycle, thereby providing a new duty factor D_(LR), andadds a predetermined duty factor ΔD2 to the duty factor D24 of thesolenoid valve 47, thereby providing a new duty cycle D24. Then, the TCU16 delivers an output signal such that the solenoid valves 47 and 48 areoperated with the new duty factors D_(LR) and D24, respectively (stepS30). The added predetermined duty factor ΔD2 is set to a value suchthat the duty factor D24 of the solenoid valve 48 increases at apredetermined rate (e.g., 15% per second) (see the transition of theduty factor D24 from time t2 to t3 in FIG. 13(c)).

Subsequently, the program proceeds to step S32, whereupon the TCU 16calculates an actual slipping rate N_(SR) according to equation (10),and compares the calculated value with a predetermined discriminationvalue ΔN_(SR1) (e.g., 10 rpm). There is a relation:

    N.sub.SR =Nt-Ntc1,                                         (10)

where Ntc1 is a calculated turbine speed for the first gear ratio mode,which is obtained by multiplying the transfer drive gear speed No,detected by means of the No sensor 17, by a predetermined number.

If the actual slipping rate N_(SR) is smaller than the predetermineddiscrimination value ΔN_(SR1) (N_(SR) <ΔN_(SR1)), the program returns tostep S28, whereupon the TCU 16 repeats the operations of steps S28 toS32. Thus, the first-speed clutch 33 on the release side is graduallydisengaged, while the second-speed clutch 34 on the connection side isyet to start engagement, although it is gradually moved from thepredetermined position just short of the engagement start position tothe engagement side. In such a situation, the turbine speed Nt graduallyincreases (in the latter half of a control section A of FIG. 13(a)) asthe first-clutch 33 is disengaged. Thus, in the control section A(between time t1 for shift signal output and time t3 for the detectionof the actual slipping rate N_(SR) attaining the predetermineddiscrimination value ΔN_(SR1) or more), the first-speed clutch 33 isgradually disengaged before the friction torque of the second-speedclutch 34 is produced. By doing this, the actual slipping rate N_(SR) isincreased toward a predetermined target slipping rate N_(SO) mentionedlater. When it is detected that the actual slipping frequency N_(SR) isnot less than the predetermined discrimination value ΔN_(SR1) (N_(SR)≧ΔN_(SR1)), the program proceeds to step S34 shown in FIG. 10.

In step S34, the TCU 16 adjusts the duty factor D24 of theconnection-side solenoid valve 48 to the initial duty factor D_(U2)calculated in step S20, and delivers an output signal such that thesolenoid valve 48 is operated with the duty factor D24. At the sametime, the TCU 15 subtracts a predetermined duty factor ΔD4 (e.g., 2 to6%) from the duty factor D_(LR) of the release-side solenoid valve 47set in the preceding duty cycle, thereby providing a new duty factorD_(LR). Using the duty factor D_(LR) as an initial value, hydrauliccontrol is started such that the actual slipping rate N_(SR) isfeedback-controlled for the predetermined target slipping rate N_(SO)(step S35). Thus, the TCU 16 waits until one duty cycle ends in stepS36, then sets the duty factor D_(LR) of the release-side solenoid valve47 for each duty cycle in the following manner, and delivers a drivingsignal such that the solenoid valve 47 is operated with the set dutyfactor D_(LR) (step S38). There is a relation:

    (D.sub.LR).sub.n =(Di).sub.n +K.sub.P1 ×e.sub.n +K.sub.D1 (e.sub.n -e.sub.n-1),                                              (11)

where e_(n) is the difference (e_(n) =N_(SO) -N_(SR)) between the actualslipping rate N_(SR) and the target slipping rate N_(SO) for the presentduty cycle, and e_(n-1) is the difference between the actual slippingrate N_(SR) and the target slipping rate N_(SO) for the preceding dutycycle. K_(P1) and K_(D1) are a proportional gain and a differentialgain, respectively, which are set to their respective predeterminedvalues. (Di)_(n) is an integral term, which is calculated as follows:

    (Di).sub.n =(Di).sub.n-1 +K.sub.I1 ×e.sub.n +D.sub.H1,(11a)

where (Di)_(n-1) is an integral term set in the preceding duty cycle,and K_(I1) is a integral gain set to a predetermined value.

D_(H1) is a correction of the turbine shaft torque set in accordancewith a variation ΔTt of the turbine shaft torque caused when the enginetorque Te is changed by accleration work during the transmission controloperation. The variation ΔTt is first calculated, and the correctionD_(H1) responsive to the variation ΔTt is then calculated as follows:

    D.sub.H1 =a6×ΔTt,                              (12)

where ΔTt, in this power-on region, is given by

    ΔTt=(Tt).sub.n -(Tt).sub.n-1.                        (13)

In a power-off region mentioned later, ΔTt is

    ΔTt=-(Tt).sub.n +(Tt).sub.n-1,                       (14)

where (Tt)_(n) and (Tt)_(n-1) are turbine shaft torques for the presentand preceding duty cycles, respectively, which are set in step S13 ofFIG. 4. In equation (12), a6 is a constant previously set in accordancewith the shift schedule. As seen frolm equations (11a) and (12), theintegral term (Di)_(n) includes the duty factor correction D_(H1) whichcan be obtained on the basis of the variation ΔTt of the turbine shafttorque. Accordingly, the duty factor D_(LR) can be corrected withoutdelay behind the change of the turbine shaft torque. Thus, the aforesaidintegral, proportional, and differential gains for the feedback controlneed not be set to great values, so that stable control can be effectedwith satisfactory follow-up performance.

Subsequently, the TCU 16 determines whether the actual slipping rateN_(SR) is not more than a negative predetermined slipping rate ΔN_(S1)(e.g., -3 to -7 rpm) (step S40). If the conclusion of step S40 is NO,the program returns to step S36, whereupon the TCU 16 repeatedlyexecutes steps S36 to S40 until the actual slipping rate N_(SR) becomesnot more than the predetermined frequency ΔN_(S1). Thereupon, the dutyfactor D_(LR) of the release-side solenoid valve 47 isfeedback-controlled so that the difference between the actual slippingrate N_(SR) and the target slipping rate N_(SO) is reduced, or that therates N_(SR) and N_(SO) are equal. On the other hand, the duty factorD24 of the connection-side solenoid valve 48 is kept constant at thelevel of the initial duty factor D_(U2). As a result, an operating oilpressure corresponding to the initial duty factor D_(U2) of the solenoidvalve 48 is supplied to the second-speed clutch 34 via the secondhydraulic control valve 46, so that the piston (not shown) of the clutch34 gradually moves to the engagement side. Thus, the clutch 34 starts toengage, so that the turbine speed Nt is motivated to lower. Since theengine 10 is in the power-on state, however, the turbine speed Nt can beprevented from lowering by setting the duty factor D_(LR) of therelease-side solenoid valve 47 to a greater value. However, when theengagement of the engagement-side clutch 34 advances so that theengaging force of the clutch 34 exceeds the relatively great value ofthe duty factor D_(LR) of the release-side solenoid valve 47, theturbine speed Nt starts to lower. At time t4 to FIG. 13(a)), the actualslipping rate N_(SR) becomes not more than the negative predeterminedslipping rate ΔN_(S1). When this is detected (YES in step S40), theprogram advances to step S42 shown in FIG. 11. Thus, hydraulic controlin a control section B (between times t3 and t4) shown in FIG. 13 isfinished.

If the actual slipping rate N_(SR) is found to be not more than thenegative predetermined slipping rate ΔN_(S1) in the control section B,step S42 of FIG. 11 is executed. If it is detected twice in twoconsecutive duty cycles, for example, in the control section A that theactual slipping rate N_(SR) is lowered to the level of the negativepredetermined slipping rate ΔN_(S1) or less by any disturbance, thehydraulic control in the control section B may be omitted. In this case,the program proceeds directly to step S42 of FIG. 11, whereuponhydraulic control in a control section C is started.

In hydraulic control operations in the control section C and controlsections D and E subsequent thereto, the duty factor D24 of theconnection-side solenoid valve 48 is feedback-controlled so that thedifference between the turbine speed changing rate ωt and thepredetemrined target turbine speed changing rate ωto is minimized. Thus,the turbine speed Nt is gradually decreased toward a calculated turbinespeed Ntc2 for the second gear ratio mode. The TCU 16 first adjusts theduty factor D_(LR) of the release-side solenoid valve 47 to apredetermined duty factor D_(LR) max, and delivers a driving signal suchthat the solenoid valve 47 is operated with the adjusted duty factorD_(LR) (step S42). The predetermined duty factor D_(LR) max is set to avalue such that the operating oil pressure supplied through the firsthydraulic control valve 44 to the first-speed clutch 33 can be kept at afixed pressure (hold pressure), and that the piston of the clutch 33 canbe held in a position corresponding to time t4 of FIG. 13(b). Until thetransmission control is practically completed thereafter (between timest4 and t8 of FIG. 13(b)), the duty factor D_(LR) of the release-sidesolenoid valve 47 is kept at the level of the predetermined duty factorD_(LR) max which provides the hold pressure for the first-speed clutch33.

Then, after the predetermined period t_(D) terminates (step S43), theprogram advances to step S44. In step S44, the target turbine speedchanging rate ωto is set as follows:

    ωto=a7×No+b7,                                  (15)

where a7 and b7 are set to predetermined values (negative values)corresponding to the control sections C, D and E. In the control sectionC immediately after the after of the feedback control, the values a7 andb7 are adjusted so that the target turbine speed changing rate ωto isset to a value such that the turbine speed Nt decreases gradually. Inthe control section D next to the control section C, the changing rateωto is adjusted so that its absolute value is greater than in thesection C. Thus, in the section D, the turbine speed Nt decreases moresharply. In the control section E during which the engagement of thesecond-speed clutch 34 finishes, the absolute value of the changing rateis reduced again to prevent the shock of the transmission control (seethe time-based transition of the turbine speed Nt of FIG. 13(a)).

Then, the TCU 16 calculates and sets the duty factor D24 of theconnection-side solenoid valve 48 according to the following equation(16), using the duty factor obtained, as an initial value, at time t4when it is detected that the actual slipping rate N_(SR) is reduced tothe level of the negative predetermined slipping rate ΔN_(S1) or less.Then, the TCU 16 delivers a driving signal such that the solenoid valve48 is operated with the set duty factor D24 (step S46). There is arelation:

    (D24).sub.n =(Di).sub.n +K.sub.P2 ×E.sub.n +K.sub.D2 (E.sub.n -E.sub.n-1),                                              (16)

where E_(n) is the difference (E_(n) =ωto-ωt) between the actual turbinespeed changing rate ωt and the target turbine speed changing rate ωtofor the present duty cycle set in step S44. Based on actual turbinespeeds (Nt)_(n) and (Nt)_(n-1) for the present and preceding dutycycles, the actual changing rate ωt is obtained as follows:

    (ωt).sub.n =(Nt).sub.n -(Nt).sub.n-1.                (17)

E_(n-1) is the difference between the actual turbine speed changing rateωt and the target turbine speed changing rate ωto for the preceding dutycycle. K_(P2) and K_(D2) are a proportional gain and a differentialgain, respectively, which are set to their respective predeterminedvalues. (Di)_(n) is an integral term, which is calculated as follows:

    (Di).sub.n =(Di).sub.n-1 +K.sub.I2 ×E.sub.n +D.sub.H1 +D.sub.H2,(18)

where (Di)_(n-1) is an integral term set in the preceding duty cycle,and K_(I2) is a integral gain set to a predetermined value.

D_(H1) is a correction of the turbine shaft torque set in accordancewith a variation ΔTt of the turbine shaft torque caused when the enginetorque Te is changed by acceleration work during the transmissioncontrol operation. The correction D_(H1) is obtained according to thesame equations as equations (12) to (14).

D_(H2) is a corrected duty factor for the change of the target turbinespeed changing rate, which is used only when the control section changesfrom C to D or from D to E. This value is obtained as follows:

    D.sub.H2 =α×Δωto,                  (19)

    Δ ωto=(ωto).sub.n -(ωto).sub.n-1,  (20)

where (ωto)_(n) is a target turbine speed changing rate to be used forthe present duty cycle and the subsequent duty cycles, and (ωto)_(n-1)is a target turbine speed changing rate having so far been used for thepreceding duty cycles. In equation (19), moreover, α is a constant setaccording to the shift schedule.

Like the integral term of the duty factor D_(LR) of the release-sidesolenoid valve 47 calculated in the control section B, the integral term(Di)_(n) of the duty factor D24 calculated for each duty cycle iscorrected on the basis of the duty factor correction D_(H1), i.e., thevariation ΔTt of the turbine shaft torque. When the control section ischanged, the integral term (Di)_(n) is corrected in accordance with thevariation Δ ωto of the target turbine speed changing rate. Accordingly,the duty factor D24 can be corrected without delay behind the changes ofthe turbine shaft torque and the target turbine speed changing rate.Thus, the aforesaid integral, proportional, and differential gains forthe feedback control need not be set to great values, so thathunting-free stable control can be effected with satisfactory follow-upperformance.

After calculating the duty factor D24 and delivering the driving signalin step S46, the TCU 16 proceeds to step S48, and determines whether theturbine speed Nt has attained a predetermined speed Ntc20 which ishigher than the calculated turbine speed Ntc2 for the second gear ratiomode by ΔNtc2 (e.g., 80 to 120 rpm). If the conclusion of step S46 isNO, the program returns to step S43, and the operations of steps S43 toS48 are repeated.

At the time immediately after the control section C is entered, theengagement of the connection-side clutch 34 is just started. Therefore,the shock of the transmission control at the start of the engagement canbe avoided by reducing the turbine speed Nt at the aforesaid targetturbine speed changing rate ωto. When the turbine speed Nt is reduced tothe level of the product of the transfer drive gear speed No and apredetermined coefficient (e.g., 2.8), the TCU 16 concludes that thecontrol section C is left, that is, the control section D is entered,and changes the absolute value of the target changing rate ωto in stepS44 into a greater value (at time t5 of FIG. 13(a)).

If the absolute value of the target turbine speed changing rate ωto ischanged into the greater value, the duty factor D24 of theconnection-side solenoid valve 48 is adjusted (during the period betweentimes t5 and t6 of FIG. 13(c)) to a value greater than the value set inthe control section C. Thus, the turbine speed Nt is quickly loweredsubstantially at the target changing rate ωto. The greater the absolutevalue of the target value ωto, the higher the transmission-controlresponsiveness will be.

Subsequently, when the turbine speed Nt is further reduced to the levelof the product of the transfer drive gear speed No and anotherpredetermined coefficient (e.g., 2.2), that is, when the piston of thesecond-speed clutlch 34 is gradually moved close to the engagement endposition, the TCU 16 concludes that the control section D is left, thatis, the control section E is entered, and changes the absolute value ofthe target turbine speed changing rate ωto in step S44 into a valuesmaller than the value set in the control section D (at time t6 of FIG.13(a)). If the absolute value of the target turbine speed changing rateωto is changed into the smaller value, the duty factor D24 of theconnection-side solenoid valve 48 is adjusted (during the period betweentimes t6 and t7 of FIG. 13(c)) to a value smaller than the value set inthe control section D. Thus, the turbine speed Nt is slowly loweredsubstantially at the target changing rate ωto. As a result, therelease-side clutch 33 is disengaged completely, so that thetransmission-control shock can be avoided near the time when theengagement of the connection-side clutch 34 is completed.

If the conclusion of step S48 is YES, that is, when the turbine speed Ntattains the level of the predetermined speed Ntc20 just higher than thecalculated turbine speed Ntc2 for the second gear ratio mode (at time t7of FIG. 13(c)), the TCU 16 sets a predetermined period T_(SF) (e.g., 0.5sec) in the aforesaid timer (step S50), and waits until the periodT_(SF) terminates (step S51). By doing this, the TCU 16 can securelycomplete the engagement of the connection-side clutch 34.

When the predetermined period T_(SF) terminates so that the conclusionof step S51 is YES, the TCU 16 sets both the duty factors D_(LR) and D24of the release- and connection-side solenoid valves 47 and 48 to 100%,and delivers a driving signal such that the solenoid valves 47 and 48are operated with the duty factors D_(LR) and D24 (at time t8 of FIGS.13(b) and 13(c)). Thus, the hydraulic transmission control for thepower-on up-shift from the first gear ratio mode to the second iscompleted.

Hydraulic Control for Power-On Down-Shift

FIGS. 14 to 16 are flow charts showing processes of hydraulictransmission control for a power-on down-shift mode. Referring to FIG.17, the processes of hydraulic control will be described in connectionwith the case of shift-down operation from the second gear ratio mode tothe first, by way of example.

In response to a shift signal for a power-on down-shift from the secondgear ratio mode to the first, the TCU 16 first calculates the respectiveinitial duty factors D_(d1) and D_(d2) of the solenoid valves 47 and 48according to the following equations (21) and (22) which resembleequations (8) and (9), respectively (step S60).

    D.sub.d1 =a8×|Tt|+c8,              (21)

    D.sub.d2 =a9×|Tt|+c9,              (22)

where a8, c8, a9, and c9 are constants applied to the case of theshift-down operation from the second gear ratio mode to the first.

Then, the TCU 16 adjusts the duty factor D24 of the release-sidesolenoid valve 48 to the initial duty factor D_(d1) set in step S60, anddelivers an output signal such that the solenoid valve 48 is operatedwith the duty factor D24. Thereupon, the second-speed clutch 34, for useas the release-side frictional engagement element, starts to be suppliedwith an initial oil pressure corresponding to the initial duty factorD_(d1), so that the piston (not shown) of the clutch 34 is retreatedtoward a position just short of the position where the clutch 34 slips(step S62; time t10 of FIG. 17(b)). Meanwhile, the TCU 16 sets the dutyfactor D_(LR) of the connection-side solenoid valve 47 to 0%, anddelivers an output signal such that the solenoid valve 47 is operatedwith the duty factor D_(LR). Thus, the normally-open solenoid valve 47is fully opened. Thereupon, the piston of the first-speed clutch 33, foruse as the connection-side frictional engagement element, is movedtoward a position just before the clutlch 33 starts to engage (at timet10 of FIG. 17(c)), and an initial pressure supply period T_(S2) is setin the timer (step S64). If the normally-open solenoid valve 47 isdriven to supply the connection-side clutch 33 with the opening oilpressure with 0-% duty factor all during the period T_(S2), the pistonof the clutch 33 can be advanced to the predetermined position justshort of the engagement start position.

The TCU 16 determines whether the initial pressure supply period T_(S2),set in step S64, is over (step S66). If the period T_(S2) is not overyet, the TCU 16 waits until the period T_(S2) terminates, repeatedlyexecuting the operation of step S66.

If the conclusion of step S66 is YES, that is, when the first-speedclutch 33 advances to the predetermined position just short of theengagement position after the termination of the initial pressure supplyperiod T_(S2), the program proceeds to step S68 of FIG. 15. In step S68,the TCU 16 adjusts the duty factor D_(LR) of the connection-sidesolenoid value 47 to the predetermined value D_(LR) max for the holdpressure, and then delivers a driving signal such that the valve 47 isoperated with the duty factor D_(LR) (at time t11 of FIG. 17(c)). Theduty factor D_(LR) of the connection-side solenoid valve 47 is kept atthe level of the predetermined duty factor D_(LR) max for the holdpressure on the first-speed clutch 33 until the turbine speed Ntattains, thereafter, the calculated turbine speed Ntc1 for the firstgear ratio mode (during the period between times t11 and t15 of FIG.17(a)).

Meanwhile, the piston of the release-side clutch 34 gradually moves tothe side for disengagement, thereby reducing the friction torque of theclutch 34, so that the turbine speed Nt starts of increase gradually.Then, the TCU 16 determines whether the turbine speed Nt is increasedabove a first predetermined discrimination value (e.g., 1.5×No) (stepS70). If the discrimination value (1.5×No) is not exceeded, the TCU 16waits until the predetermined speed value is exceeded, repeating thediscrimination of step S70.

If the turbine speed Nt exceeds the speed (1.5×No) (at time t12 of FIG.17(a)), then it indicates that hydraulic transmission control in acontrol section A of FIG. 17 is finished, and that a control section Bis entered. In step S71 subsequent to step S70, the TCU 16 waits untilone duty cycle terminates. Thereafter, the TCU 16 starts a hydrauliccontrol operation such that the turbine speed Nt is increased toward thecalculated turbine speed Ntc1 for the first gear ratio mode whileadjusting the turbine speed changing ratio ωt by feedback control. Thus,in hydraulic control operations in the control section B and controlsections C and D subsequent thereto, the duty factor D24 of therelease-side solenoid valve 48 is feedback-controlled so as to take avalue such that the difference between the changing rate ωt and thepredetermined target turbine speed changing rate ωto is minimized. Also,the turbine speed Nt is gradually increased toward the level of thecalculated turbine speed Ntc1 for the first gear ratio mode.

In step S72, the TCU 16 first sets the target turbine speed changingrate ωto as follows:

    ωto=a10×No+b10,                                (23)

where a10 and b10 are constants which are set to predetermined values(positive values) corresponding to the control sections B, C and D. Inthe control section B immediately after the start of the feedbackcontrol, the values a10 and b10 are adjusted so that the target turbinespeed changing rate ωto is set to a value such that the turbine speed Ntincreases gradually. In the control section C next to the controlsection B, the changing rate is adjusted to a greater value than in thesection B. Thus, in the section C, the turbine speed Nt increases moresharply. In the control section D during which turbine speed Ntapproaches the calculated turbine speed Ntc1 for the first gear ratiomode, the changing rate is reduced again to prevent an overshoot of theturbine speed Nt (see the time-based transition of the turbine speed Ntof FIG. 17(a)).

Then, the TCU 16 calculates and sets the duty factor D24 of therelease-side solenoid valve 48 according to the same equations asequations (16) and (18), using the duty factor obtained, as an initialvalue, at time t12 when the turbine speed Nt exceeds the speed (1.5×No).Then, the TCU 16 delivers a driving signal such that the solenoid valve48 is operated with the set duty factor D24 (step S74). The integral,proportional, and differential gains K_(I2), K_(P2), and K_(D2) inequations (16)and (18) are set to their respective optimum values forthe shift schedule in the power-on-down-shift mode.

After calculating the duty factor D24 and delivering the driving signalin step S74, the TCU 16 proceeds to step S76, and determines whether theturbine speed Nt has attained the level of the calculated turbine speedNtc1 for the first gear ratio mode. If the conclusion of step S76 is NO,the program returns to step S71, and the operations of steps S71 to S76are repeated.

At the immediately after the control section B is entered, thedisengagement of the release-side clutch 34 is just started. Therefore,an overshoot of the turbine speed Nt can be avoided by increasing theturbine speed Nt at the aforesaid target turbine speed changing rateωto. When the turbine speed Nt is increased to the level of the productof the transfer drive gear speed No and a predetermined coefficient(e.g., 1.7), the TCU 16 concludes that the control section B is left,that is, the control section C is entered, and changes the targetturbine speed changing rate ωto into a greater value in step S72 (attime t13 of FIG. 17(a)).

If the target turbine speed changing rate ωto is changed into thegreater value, the duty factor D24 of the release-side solenoid value 48is adjusted (during the period between times t13 and t14 of FIG. 17(b))to a value smaller than the value set in the control section B. Thus,the turbine speed Nt is quickly increased substantially at the targetchanging rate ωto. The greater the target changing rate ωto, the higherthe transmission-control responsiveness will be.

Subsequently, when the turbine speed Nt is further increased to thelevel of the product of the transfer drive gear speed No and anotherpredetermined coefficient (e.g., 2.4), that is, when the second-speedclutch 34 is gradually disengaged so that the turbine speed Ntapproaches the calculated turbine speed Ntc1 for the first gear ratiomode, the TCU 16 concludes that the control section C is left, that is,the control section D is entered, and changes the target turbine speedchanging rate ωto in step S72 into a value smaller than the value set inthe control section C (at time t14 of FIG. 17(a)). If the target turbinespeed changing rate ωto is changed into the smaller value, the dutyfactor D24 of the release-side solenoid valve 48 is adjusted (during theperiod between times t14 and t15 of FIG. 17(b)) to a value greater thanthe value set in the control section C. Thus, the turbine speed Nt isslowly increased substantially at the target changing rate ωto, andtherefore, can be prevented from substantially overshooting thecalculated turbine speed Ntc1 for the first gear ratio mode.

If the conclusion of step S76 is YES, and when the turbine speed Nt isfound to be on the same level as the calculated turbine speed Ntc1 forthe first gear ratio mode (at time t15 of FIG. 17(a)), the hydrauliccontrol operation in the control section D is finished, and a hydrauliccontrol operatoin in a control section E is started. In the hydrauliccontrol operation in the control section E, the duty factor D24 of therelease-side solenoid valve 48 is feedback-controlled so that thedifference between the actual slipping rate N_(SR) and the targetslipping rate N_(SO) (e.g., 20 rpm) is minimized, and the engagement ofthe first-speed clutch 33 on the connection side is graduallystrengthened in the meantime. Thus, in step S78, the TCU 16 adjusts theduty factor D_(LR) of the connection-side solenoid valve 47 to theinitial duty factor D_(d2), which, set in step S60, is smaller than theduty factor D_(LR) max, and delivers a driving signal such that thesolenoid valve 47 is operated with the duty factor D_(LR) (at time t15of FIG. 17(c)). Thereupon, the piston of the first-speed clutch 33 onthe connectio side starts to move gradually to the engagement side.

Subsequently, the TCU 16 waits until the predetermined period t_(D)terminates in step S79, then calculates the duty factor D24 fo therelease-side solenoid valve 48 for each duty cycle, according to thefollowing equations (24) and (24a) which resemble equations (11) and(11a), respectively, and delivers a driving signal such that thesolenoid valve 48 is operated with the duty factor D24 (step S80). Thereare relations:

    (D24).sub.n =(Di).sub.n +K.sub.P1 ×e.sub.n +K.sub.D1 (e.sub.n -e.sub.n-1),                                              (24)

    (Di).sub.n =(Di).sub.n-1 +K.sub.I1 ×e.sub.n +D.sub.H1,(24a)

where (Di)_(n-1) is an integral term set in a preceding duty cycle. Usedas the initial value is a duty factor which is set immediately beforetime t15 when it is detected that the calculated turbine speed Ntc1 forthe first gear ratio mode is exceeded by the turbine speed Nt. K_(I1),K_(P1), and K_(D1) are integral, proportional, and differential gains,which are set to their respective optimum values for the power-ondown-shift mode. In equation (24), e_(n) is the difference (e_(n)=N_(SO) -N_(SR)) between the actual slipping rate N_(SR) and the targetslipping rate N_(SO) for the present duty cycle, and e_(n-1) is thedifference between the actual slipping rate N_(SR) and the targetslipping rate N_(SO) for the preceding duty cycle.

D_(H1) is a correction of the turbine shaft torque set in accordancewith a variation ΔTt of the turbine shaft torque caused when the enginetorque Te is changed by acceleration work during the transmissioncontrol operation. The value D_(H1) is calculated according to equations(12) to (14).

Then, in steps S82 to S85, the TCU 16 determines whether it is detectedin two consecutive duty cycles that the absolute value of the actualslipping rate N_(SR) is smaller than a predetermined value (e.g., 5rpm). More specifically, in step S82, the TCU 16 determines whether theabsolute value of the actual slipping rate N_(SR) is smaller than thepredetermined value (5 rpm). If the conclusion of step S82 is NO, theTCU 16 resets a flag value FLG to 0 (step S83). Thereupon, the programreturns to step S79, and the operations of steps S79 to S82 arerepeated. If the friction torque of the connection-side clutch 33 issmall, and as long as the reduction of the friction torque of the clutch34, i.e., the degree of disengagement thereof, is made greater than theincrease of the friction torque of the clutch 33 by the feedback controlso that the torque to increase the turbine speed Nt by means of theengine 10 in the power-on state is prevailing, the turbine speed Nt canbe kept at a level higher than the calculated turbine speed Ntc1 for thefirst gear ratio mode by the target slipping rate N_(SO). If thefriction torque of the clutch 33 becomes higher, however, the turbinespeed Nt gradually lowers so that the conclusion of step S82 turns toYES, whereupon the operation of step S84 is executed.

In step S84, the TCU determines whether the flag value FLG is equalto 1. If the turbine speed Nt lowers so that the conclusiokn of step S82becomes YES for the first time, the conclusion of step S84 is NO. Inthis case, the flag value FLG is adjusted to 1 in step S85, whereuponthe program returns to step S79, and the operations of steps S79 and S80are executed. If it is detected twice, for two consecutive duty cycles,again in step S82 that the absolute value of the actual slipping rateN_(SR) is smaller than the predetermined value (5 rpm) (at time t16 ofFIG. 17(a)), the conclusion of step S84 is YES. Thus, hydraulic controlin the control section E is finished, and the operation of step S87 isexecuted.

In step S87, the TCU 16 sets both the duty factors D_(LR) and D24 of theconnection- and release-side solenoid valves 47 and 48 to 0%, anddelivers no driving signal to the valves 47 and 48. Thus, thedisengagement of the second-speed clutch 34 and the engagement offirst-speed clutch 33 are finished, whereupon the hydraulic transmissioncontrol for the power-on down-shift from the second gear ratio mode tothe first is completed.

Hydraulic Control for Power-Off Up-Shift

FIGS. 18 to 20 are flow charts showing processes of hydraulictransmission control for a power-off up-shift mode. Referring to FIG.21, the processes of hydraulic control will be described in connectionwith the case of shift-up operation from the first gear ratio mode tothe second, by way of example.

In response to a shift signal for a power-off up-shift from the firstgear ratio mode to the second, the TCU 16 first calculates the initialduty factor D_(U2) of the connection-side solenoid valves 48 accordingto the same equation as equation (9) (step S90).

Then, the TCU 16 adjusts the duty factor D_(LR) of the release-sidesolenoid valve 47 to the predetermined duty factor D_(LR) max for thehold pressure, and delivers an output signal such that the solenoidvalve 47 is operated with the duty factor D_(LR). Thereupon, the piston(not shown) of the first-speed clutch 33, for use as the release-sidefrictional engagement element, is retreated toward a stand-by posiitonwhere the clutch 33 slips entirely and the engagement can be resumed atonce (step S92); time 21 of FIG. 21(b)). When the engine 10 is in apower-off operation mode, there is no possibility of the turbine speedNt overshooting even if the release-side clutch 33 is disengagedimmediately after the delivery of the shift signal. On the contrary, atransmission control shock may possibly be caused unless the clutch 33is released without delay. Meanwhile, the TCU 16 sets the duty factorD24 of the connection-side solenoid valve 48 to 100%, and delivers anoutput signal such that the solenoid valve 48 is operated with the dutyfactor D24, i.e., a driving signal to fully open the solenoid valve 48.Thereupon, a piston of the second-speed clutch 34, for use as aconnection-side frictional engagement element, is advanced toward aposition just before the clutch 34 starts to engage (at time t21 of FIG.21(c)), and the initial pressure supply period T_(S1) is set in thetimer (step S93).

The TCU 16 determines whether the initial pressure supply period T_(S1),set in step S93, is over (step S95). If the period T_(S1) is not over,the step S95 is repeated until the period T_(S1) terminates.

If the conclusion of step S95 is YES, that is, when the second-speedclutch 34 advances to the predetermined position just short of theengagement position after the termination of the initial pressure supplyperiod T_(S1), the program proceeds to stop S96. In step S96, the TCU 16adjusts the duty factor D24 of the connection-side solenoid value 48 tothe initial duty factor D_(U2) calculated in step S90, and then deliversa driving signal such that the valve 48 is operated with the duty factorD24 (at time t22 of FIG. 21(c)). When the predetermined period t_(D) forone duty cycle terminates (step S98), the TCU adds a predetermined dutyfactor ΔD5 to the duty factor D24 of the solenoid valve 48 set in thepreceding duty cycle, thereby providing a new duty factor D24, anddelivers an output signal such that the solenoid valve 48 is operatedwith the new duty factor D24 (step S99). The added predetermined dutyfactor ΔD5 is set to a value such that the duty factor D24 of thesolenoid valve 48 increases at a predetermined rate (e.g., 14 to 17% persecond) (see the transition of the duty factor D24 from time t22 to t23in FIG. 21(c)).

Subsequently, the program proceeds to step S100, whereupon the TCU 16calculates the actual slipping rate N_(SR) according to equation (10),and compares the calculated value with a negative predetermineddiscrimination value ΔN_(SR2) (e.g., -8 to -12 rpm).

If the actual slipping rate N_(SR) is greater than the predetermineddiscrimination value ΔN_(SR2) (N_(SR) >ΔN_(SR2)), the program returns tostep S98, whereupon the TCU 16 repeats the operations of steps S98 toS100, thereby gradually increasing the duty factor D24 of the solenoidvalve 48. Thus, the connection-side clutch 34 starts to engage, so thatthe friction torque of the clutch 34 increases gradually. Thereupon, theturbine speed Nt lowers gradually, so that the conclusion of step S100becomes YES. Then the program proceeds to step S102 of FIG. 19,whereupon hydraulic control in a control section A is finished, andhydraulic control in a control section B is started.

In the hydraulic control operations in the control section B and controlsections C and D subsequent thereto, the duty factor D24 of theconnection-side solenoid valve 48 is feedback-controlled so that thedifference between the turbine speed changing rate ωt and thepredetermined target turbine speed changing rate ωto is minimized. Thus,the turbine speed Nt is gradually decreased toward the calculatedturbine speed Ntc2 for the second gear ratio mode.

After waiting until one duty cycle or the predetermined period t_(D)terminates in step S102, the program first sets the target turbine speedchanging rate ωto to predetermined values previously storedcorresponding to the control sections B to D. In the control section Bimmediately after the start of the feedback control, the target turbinespeed changing rate ωto is set to a value such that the turbine speed Ntdecreases gradually. In the control section C next to the controlsection B, the changing rate ωto is adjusted so that its absolute valueis greater than in the section B. Thus, in the section C, the turbinespeed Nt decreases more sharply. In the control section D during whichthe engagement of the second-speed clutch 34 substantially finishes sothat the turbine speed Nt approaches the turbine speed Ntc2 for thesecond gear ratio mode, the absolute value of the changing rate isreduced again to prevent the shock of the transmission control (see thetime-based transition of the turbine speed Nt of FIG. 21(a)).

Then, the TCU 16 calculates and sets the duty factor D24 of theconnection-side solenoid valve 48 according to equations (16) and (18),using the duty factor obtained, as an initial value, at time t23 when itis detected that the actual slipping rate N_(SR) is reduced to the levelof the negative predetermined slipping rate ΔN_(S2) (e.g., -8 to -12rpm) or less. Then, the TCU 16 delivers a driving signal such that thesolenoid valve 48 is operated with the set duty factor D24 (step S106).The integral, proportional, and differential gains K₁₂, K_(p2), andK_(D2) in equations (16) and (18) are set to their respective optimumvalues for the shift schedule in the power-off up-shift mode.

After calculating the duty factor D24 and delivering the driving signalin step S106, the TCU 16 proceeds to step S107, and determines whetherthe turbine speed Nt is lowered to the predetermined speed Ntc20 whichis higher than the calculated turbine speed Ntc2 for the second gearratio mode by ΔNtc2 (e.g., 80 to 120 rpm). If the conclusion of stepS107 is NO, the program returns to step S102, and the operations ofsteps S102 to S107 are repeated.

At the time immediately after the control section B is entered, theengagement of the connection-side clutch 34 is just started. Therefore,the shock of the transmission control at the start of the engagement canbe avoided by reducing the turbine speed Nt at the aforesaid targetturbine speed changing rate ωto. When the turbine speed Nt is reduced tothe level of the product of the transfer drive gear speed No and apredetermined coefficient (e.g., 2.8), the TCU 16 concludes that thecontrol section B is left, that is, the control section C is entered,and changes the absolute value of the target changing rate ωto in stepS104 into a value greater than the value used in the control section C(at time t24 of FIG. 21(a)).

If the absolute value of the target turbine speed changing rate ωto ischanged into the greater value, the duty factor D24 of theconnection-side solenoid valve 48 is adjusted (during the period betweentimes t24 and t25 of FIG. 21(c)) to a value greater than the value setin the control section B. Thus, the turbine speed Nt is quickly loweredsubstantially at the target changing rate ωto set in this manner. Thegreater the absolute value of the target value ωto, the higher thetransmission-control responsiveness will be.

Subsequently, when the turbine speed Nt is further reduced to the levelof the product of the transfer drive gear speed No and anotherpredetermined coefficient (e.g., 2.2), that is, when the piston of thesecond-speed clutch 34 is gradually moved close to the engagement endposition, the TCU 16 concludes that the control section C is left, thatis, the control section D is entered, and changes the absolute value ofthe target turbine speed changing rate ωto in step S104 into a valuesmaller than the value set in the control section C (at time t25 of FIG.21(a)). If the absolute value of the target turbine speed changing rateωto is changed into the smaller value, the duty factor D24 of theconnection-side solenoid valve 48 is adjusted (during the period betweentimes t25 and t26 of FIG. 21(c)) to a value smaller than the value setin the control section C. Thus, the turbine speed Nt is slowly loweredat a rate substantially equal to the target changing rate ωto. As aresult, the turbine speed Nt is changed so smoothly into the calculatedturbine speed Ntc2 for the second gear ratio mode that thetransmission-control shock can be avoided near the time when theengagement of the connection-side clutch 34 is completed.

If the conclusion of step S107 is YES, that is, when the turbine speedNt attains the level of the predetermined speed Ntc20 just higher thanthe calculated turbine speed Ntc2 for the second gear ratio mode (attime t26 of FIG. 21(c)), the TCU 16 sets a predetermined period T_(SF)(e.g., 0.5 sec) in the aforesaid timer (step S109), and waits until theperiod T_(SF) terminates (step S110). By doing this, the TCU 16 cansecurely complete the engagement of the connection-side clutch 34.

When the predetermined period T_(SF) terminates so that the conclusionof step S110 is YES, the program proceeds to step S112, and the TCU 16sets both the duty factors D_(LR) and D24 of the release- andconnection-side solenoid valves 47 and 48 to 100%, and delivers adriving signal such that the solenoid valves 47 and 48 are operated withthe duty factors D_(LR) and D24 (at time t27 of FIGS. 21(b) and 21(c)).Thus, the hydraulic transmission control for the power-off up-shift fromthe first gear ratio mode to the second is completed.

Hydraulic Control for Power-Off Down-Shift

FIGS. 22 to 24 are flow charts showing processes of hydraulictransmission control for a power-off down-shift mode. Referring to FIG.25, the processes of hydraulic control will be described in connectionwith the case of shift-down operation from the second gear ratio mode tothe first, by way of example.

In response to a shift signal for a power-off down-shift from the secondgear ratio mode to the first, the TCU 16 first calculates the respectiveinitial duty factors D_(d1) and D_(d2) of the solenoid valves 47 and 48according to equations (21) and (22) (step S114). The values a8, c8, a9,and c9 in equations (21) and (22) are set to their respective optimumvalues for the power-off down-shift from the second speed to the first.

Then, the TCU 16 adjusts the duty factor D24 of the release-sidesolenoid valve 48 to the initial duty factor D_(d1) set in step S114,and delivers an output signal such that the solenoid valve 48 isoperated with the duty factor D24. Thereupon, the piston (not shown) ofthe second-speed clutch 34, for use as the release-side frictionalengagement element, is retreated toward the position just before wherethe clutch 34 slips (step S115; time t31 of FIG. 25(b)). Meanwhile, theTCU 16 sets the duty factor D_(LR) of the connection-side solenoid valve47 to 0%, and delivers an output signal such that the solenoid valve 47is operated with the duty factor D_(LR). Thereupon, the piston of thefirst-speed clutch 33, for use as the connection-side frictionalengagement element, is moved toward a position just short of theposition where the clutch 33 starts to engage (at time t31 of FIG.25(c)), and the initial pressure supply period T_(S2) is set in thetimer (step S116).

The TCU 16 waits until the predetermined period t_(D), i.e., one dutycycle (28.6 msec), terminates (step S118), and then subtracts apredetermined duty factor ΔD6 from the duty factor D24 set in thepreceding duty cycle, thus providing a new duty factor D24. Then, theTCU 16 delivers an output signal such that the solenoid valve 48 isoperated with the duty cycle D24 (step S120). The subtractedpredetermined duty factor ΔD6 is set to a value such that the dutyfactor D24 of the solenoid value 48 decreases at a predetermined rate(e.g., 8 to 12% per second) (see the transition of the duty factor D24from time t31 to t33 in FIG. 25(b)). The TCU 16 determines whether theinitial pressure supply period T_(S2), set in step S116, is over (stepS122). If the period T_(S2) is not over, the program returns to stepS118, and steps S118 to S122 are repeated. As a result, the duty factorD24 of the solenoid valve 48 is gradually reduced so that therelease-side clutch 34 gradually moves toward the engagement startposition.

If the conclusion of step S122 is YES, that is, when the first-speedclutch 33 advances to the predetermined position just short of theengagement start position after the termination of the initial pressuresupply period T_(S2), the program proceeds to step S124 of FIG. 23. Instep S124, the TCU 16 adjusts the duty factor D_(LR) of the solenoidvalue 47 to the initial duty factor D_(d2) calculated in step S114, andthen delivers a driving signal such that the valve 47 is operated withthe duty factor D_(LR) (at time t32 of FIG. 25(c)). Thereupon, thepiston of the connection-side clutch 33 continues to move graduallytoward the engagement start position. The duty factor D_(LR) of thesolenoid valve 47 is kept at the level of the initial duty factor D_(d2)until a control section C (mentioned later) is entered (at time t34 ofFIG. 25(c)).

Subsequently, when the predetermined period t_(D) for one duty cycleterminates (step S125), the TCU 16 continues (step S126) the calculationof the new duty factor D24 and delivery of the output signal for thevalve operation therewith in the same manner as in step S120. Then, theprogram proceeds to step S128, whereupon the TCU 16 calculates an actualslipping rate N_(SR) according to equation (25), and compares thecalculated value with a negative predetermined discrimination valueΔN_(SR2) (e.g., -8 to -12 rpm). There is a relation:

    N.sub.SR =Nt-Ntc2,                                         (25)

where Ntc2 is a calculated turbine speed for the second gear ratio mode,which is obtained by multiplying the transfer drive gear speed No by apredetermined number.

If the actual slipping rate N_(SR) is greater than the negativepredetermined discrimination value ΔN_(SR2) (N_(SR) >ΔN_(SR2)), theprogram returns to step S125, whereupon the TCU 16 repeats theoperations of steps S125 to S128. Thus, the second-speed clutch 34 onthe release side is gradually disengaged. If the first-speed clutch 33on the connection side, at this time, is yet to start engagement, theturbine speed Nt gradually decreases in the latter half of a controlsection A of FIG. 25(a) (between time t31 for shift signal output andtime t33 for the detection of the actual slipping rate N_(SR) attainingthe predetermined discrimination value ΔN_(SR2) or less). When it isdetected that the actual slipping rate N_(SR) is not more than thepredetermined discrimination value ΔN_(SR2) (N_(SR) <ΔN_(SR2)), theprogram proceeds to step S130.

In step S130, the TCU 16 adds a predetermined duty factor ΔD7 (e.g., 2to 6%) to the duty factor D24 of the release-side solenoid valve 48 setin the preceding duty cycle, thereby providing a new duty factor D24.Using this duty factor D24 as an initial value, the TCU 16 startsfeedback control such that the difference e_(n) (=N_(S1) -N_(SR))between the actual slipping rate N_(SR) and a predetermined targetslipping rate N_(S1) (e.g., -20 rpm) is minimized. When theconnection-side clutch 33 is yet to start engagement, the turbine speedNt tends to lower, due to reduction of the friction torque, if the dutyfactor D24 of the release-side clutch 34 is set to a smaller value. Ifthe duty factor D24 is set to a greater value, on the other hand, thefriction torque increases, so that the turbine speed Nt tends toincrease. Thus, the turbine speed Nt can be kept at a predeterminedvalue through the feedback control of the duty factor D24.

Thereupon, the TCU 16 waits until one duty cycle ends in step S132, andthen sets the duty factor D24 of the release-side solenoid valve 48 foreach duty cycle according to equations (24) and (24a) (step S134). Theintegral, proportional, and differential gains K_(I1), K_(P1), andK_(D1) in equations (24) and (24a) are set to their respective optimumvalues for the power-off down-shift mode.

Subsequently, the TCU 16 determines whether the actual slipping rateN_(SR) is not less than a predetermined slipping rate ΔN_(S2) (e.g., 3to 8 rpm) (step S135). If the conclusion of step S135 is NO, the programreturns to step S132, whereupon the TCU 16 repeatedly executes stepsS132 to S135 until the actual rate N_(SR) becomes not less than thepredetermined rate ΔN_(S2). Thereupon, the duty factor D24 of therelease-side solenoid valve 48 is feedback-controlled so that thedifference between the actual slipping rate N_(SR) and the targetslipping rate N_(S1) is reduced, or that the rates N_(SR) and N_(S1) areequal. On the other hand, the duty factor D_(LR) of the connection-sidesolenoid valve 47 is kept constant at the level of the initial dutyfactor D_(d2). As a result, an operating oil pressure corresponding tothe initial duty factor D_(d2) of the solenoid valve 47 is supplied tothe first-speed clutch 33 via the first hydraulic control valve 44, sothat the clutch 33 starts to engage, and the piston (not shown) of theclutch 33 gradually moves to its engagement end position. As the pistonof the clutch 33 moves in this manner, the turbine speed Nt starts toincrease. The duty factor D24 of the solenoid valve 48 is set to asmaller value such that the increase of the turbine speed Nt iscanceled, so that the value of the duty factor D24 diminishes gradually.Although the duty factor D24 of the release-side solenoid valve 48 isset to the smaller value, the turbine speed Nt increases due to anincrease of the engaging force of the engagement-side clutch 33. At timet34 of FIG. 25(a), therefore, the actual slipping rate N_(SR) becomesnot less than the predetermined slipping rate ΔN_(S2). When the TCU 16detects this (YES in step S135), the program advances to step S136 shownin FIG. 24. Thus, hydraulic control in a control section B (betweentimes t33 and t34) shown in FIG. 25 is finished.

If the actual slipping rate N_(SR) is found to be not less than thepredetermined slipping rate ΔN_(S2) in the control section B, step S136of FIG. 24 is executed. If it is detected twice in two consecutive dutycycles, for example, in the control section A that the actual slippingrate N_(SR) is increased to the level of the predetermined slipping rateΔN_(S2) or more by any disturbance, the hydraulic control in the controlsection B may be omitted. In this case, the program proceeds directly tostep S136 of FIG. 24, whereupon hydraulic control in the control sectionC is started.

In hydraulic control operations in the control section C and controlsections D and E subsequent thereto, the duty factor D_(LR) of theconnection-side solenoid valve 47 is feedback-controlled so that thedifference between the turbine speed changing rate ωt and thepredetermined target turbine speed changing rate ωto is minimized. Thus,the turbine speed Nt is gradually increased toward a calculated turbinespeed Ntc1 for the first gear ratio mode.

In step S136, the TCU 16 first adjusts the duty factor D24 of therelease-side solenoid valve 48 to a predetermined duty factor D24min forthe aforesaid hold pressure so that the hold pressure is supplied to thesecond-speed clutch 34. After waiting until the predetermined periodt_(D) terminates (step S138), the TCU 16 reads out a predetermined valuepreviously stored in the memories, corresponding to each of the controlsections C, D and E, and sets the read value as the target turbine speedchanging rate ωto (step S139). In the control section C immediatelyafter the start of the feedback control, the target turbine speedchanging rate ωto to be read in this manner is set to a small value suchthat the turbine speed Nt decreases gradually. In the control section Dnext to the control section C, the changing rate ωto is set to a greatervalue than in the section C. Thus, in the section D, the turbine speedNt decreases more sharply. In the control section E during which theengagement of the first-speed clutch 33 finishes, the changing rate isreduced again to prevent the shock of the transmission control (see thetime-based transition of the turbine speed Nt of FIG. 25(a)).

Then, the TCU 16 calculates and sets the duty factor D_(LR) of theconnection-side solenoid valve 47 according to the following equations(26) and (26a) similar to equations (16) and (18), respectively, usingthe initial duty factor D_(d2) obtained, as an initial value, at timet34 when it is detected that the actual slipping rate N_(SR) isincreased to the level of the predetermined slipping rate ΔN_(S2) ormore. Then, the TCU 16 delivers a driving signal such that the solenoidvalve 47 is operated with the set duty factor D_(LR) (step S140). Thereare relations:

    (D.sub.LR).sub.n =(Di).sub.n +K.sub.P1 ×E.sub.n +K.sub.D1 (E.sub.n -E.sub.n-1),                                              (26)

    (Di).sub.n =(Di).sub.n-1 +K.sub.I1 ×E.sub.n +D.sub.H1 +D.sub.H2,(26a)

where (Di)_(n-1) is an integral term set in the preceding duty cycle,K_(I1), K_(P1), and K_(D1) are integral, proportional, and differentialgains, which are set to their respective optimum values for thepower-off down-shift mode. In equations (26) and (26a), moreover, E_(n)is the difference between (E_(n) =ωto-ωt) between the actual turbinespeed changing rate ωt and the target turbine speed changing rate ωtofor the present duty cycle set in step S139, and E_(n-1) is thedifference between the actual turbine speed changing rate ωt and thetarget turbine speed changing rate ωto for the preceding duty cycle.

D_(H1) is a correction of the turbine shaft torque set in accordancewith a variation ΔTt of the turbine shaft torque caused when the enginetorque Te is changed by acceleration work during the transmissioncontrol operation. The correction D_(H1) is calculated according toequations (12) to (14).

D_(H2) is a corrected duty factor for the change of the target turbinespeed changing rate, which is used only when the control section changesfrom C to D or from D to E. This value is obtained from equations (19)and (20). In equation (19), the coefficient α is set to an optimum valuefor the transmission control pattern of the power-off down-shift mode.

After calculating the duty factor D_(LR) and delivering the drivingsignal in step S140, the TCU 16 proceeds to step S142, and determineswhether the turbine speed Nt has attained a predetermined speed Ntc10which is lower than the calculated turbine speed Ntc1 for the first gearratio mode by a predetermined value (e.g., 80 to 120 rpm). If theconclusion of step S142 is NO, the program returns to step S138, and theoperations of steps S138 to S142 are repeated.

At the time immediately after the control section C is entered, theengagement of the connection-side clutch 33 is just started. Therefore,the shock of the transmission control at the start of the engagement canbe avoided by increasing the turbine speed Nt at the aforesaid targetturbine speed changing rate ωto. When the turbine speed Nt is increasedto the level of the product of the transfer drive gear speed No and apredetermined coefficient (e.g., 1.7), the TCU 16 concludes that thecontrol section C is left, that is, the control section D is entered,and changes the target changing rate ωto in step S139 into a greatervalue (at time t35 of FIG. 25(a)).

If the target turbine speed changing rate ωto is changed into thegreater value, the duty factor D_(LR) of the connection-side solenoidvalve 47 is adjusted (during the period between times t35 and t36 ofFIG. 25(c)) to a value smaller than the value set in the control sectionC. Thus, the turbine speed Nt is quickly increased substantially at thetarget changing rate ωto. The greater the target value ωto, the higherthe transmission-control responsiveness will be.

Subsequently, when the turbine speed Nt is further increased to thelevel of the product of the transfer drive gear speed No and anotherpredetermined coefficient (e.g., 2.4), that is, when the piston of thefirst-speed clutch 33 is gradually moved close to the engagement endposition so that the turbine speed Nt approaches the calculated turbinespeed Ntc1 for the first gear ratio mode, the TCU 16 concludes that thecontrol section D is left, that is, the control section E is entered,and changes the target turbine speed changing rate ωto in step S139 intoa value smaller than the value set in the control section D (at time t36of FIG. 25(a)). If the target turbine speed changing rate ωto is changedinto the smaller value, the duty factor D_(LR) of the connection-sidesolenoid valve 47 is adjusted (during the period between times t36 andt37 of FIG. 25(c)) to a value greater than the value set in the controlsection D. Thus, the turbine speed Nt is slowly increased substantiallyat the target changing rate ωto. As a result, the transmission-controlshock can be avoided near the time when the engagement of theconnection-side clutch 33 is completed.

If the conclusion of step S142 is YES, that is, when the turbine speedNt attains the level of the predetermined speed Ntc10 just lower thanthe calculated turbine speed Ntc1 for the first gear ratio mode by thepredetermined value (80 to 120 rpm) (at time t37 of FIG. 25(c)), the TCU16 immediately sets both the duty factors D24 and D_(LR) of the release-and connection-side solenoid valves 48 and 47 to 0%, and delivers adriving signal such that the solenoid valves 48 and 47 are operated withthe duty factors D24 and D_(LR), respectively (at time t37 of FIGS.25(b) and 25(c)). Thus, the hydraulic transmission control for thepower-off down-shift from the second gear ratio mode to the first iscompleted.

In connection with the above embodiment, only the processes of hydrauliccontrol operation for the transmission control or shift between thefirst and second gear ratio modes have been described herein forsimplicity of illustration. It is to be understood, however, thatprocesses of hydraulic transmission control operation for any othercombination of transmission gear ratio modes, e.g., for shift betweensecond and third gear ratio modes, can be explained in like manner.

In the embodiment described above, moreover, the hydraulic clutches areused as the frictional engagement elements for transmission control.Alternatively, however, transmission control brakes may be used for thispurpose.

According the above described embodiment, furthermore, the engine torquedetection method and the hydraulic control method for a transmissioncontrol apparatus utilizing the detection method of the presentinvention is applied to an automatic transmission control apparatus witha torque converter. The driving force transmission apparatus is not,however, limited to a hydro-dynamic power transmitting device, such asthe torque converter, or a slip-type direct-coupled clutch, such as thedamper clutch 28. Various other transmission apparatuses may be used forthis purpose, provided they can substantially determinately set thetransmission torque from the rotating speed of an input or output shaft,or can externally control the transmission torque so that controlparameter values corresponding to the transmission torque can bedetected. These available transmission apparatuses include aslip-controlled electromagnetic-powder clutch, viscous clutch, etc., forexample.

The engine torque detection method according to the present invention isnot limited to the above embodiment, and may be also applied tohydraulic control for the damper clutch 28 or other slip-typedirect-coupled clutches, current control for electromagnetic-powderclutches, engine torque control, traction control, etc. In the enginetorque control, the fuel supply is adjusted so that the differencebetween the engine torque value and a target torque value based on thecar speed and the accelerator pedal stroke is minimized. In the tractioncontrol, hunting is prevented by varying the reduction rate of theengine torque when the racing acceleration is high. If the detectedengine torque is greater than a predetermined value, its reduction rateis set to a high level. If the detected value is smaller than thepredetermined level, on the other hand, the reduction rate is lowered.

What is claimed is:
 1. A method of input power on/off discrimination forselecting a transmission control logic of an automatic vehiculartransmission system having an input shaft to which an output torque ofan engine is transmitted, comprising the steps of:detecting the torquetransmitted to said input shaft; setting, as a discrimination value, avalue of torque required to maintain a rotating speed changing rate ofsaid input shaft during transmission control, at a predetermined targetvalue; comparing the detected torque of said input shaft with saiddiscrimination value; determining that said automatic vehiculartransmission system is in a power-on state if said detected torque ofsaid input shaft is higher than said discrimination value, and that saidautomatic vehicular transmission system is in a power-off state if saiddetected torque is lower than said discrimination value; and selecting asuitable transmission control logic for the automatic vehiculartransmission system based on the determined power-on or power-off stateof the transmission system.
 2. The discrimination method according toclaim 1, including determining a shift schedule based on a gear ratiomode established before a shift operation of said transmission systemand a gear ratio mode to be established, and setting said discriminationvalue to one of a number of different values in accordance with thedetermined shift schedule.
 3. The discrimination method according toclaim 2, including providing said transmission system with an outputshaft, and setting said discrimination value to a negative value whenthe transmission control is effected with a shift schedule such that theratio of the rotating speed of said output shaft to that of said inputshaft is higher.
 4. The discrimination method according to claim 2,including providing said transmission system with an output shaft, andsetting said discrimination value to a positive value when thetransmission control is effected with a shift schedule such that theratio of the rotating speed of said output shaft to that of said inputshaft is lower.
 5. The discrimination method according to claim 1,including transmitting the output torque of said engine to said inputshaft of said transmission system through a driving force transmissionapparatus capable of detecting transmission torque, detecting the speedchanging rate of said engine, detecting the transmission torque of saiddriving force transmission apparatus, adding said detected transmissiontorque and the product of said detected engine speed changing rate and apredetermined value and using the resulting sum as a calculated value ofsaid torque of said transmission system, and discriminating saidpower-on or said power-off state by using said calculated value of saidtorque.
 6. The discrimination method according to claim 5, includingproviding said driving force transmission apparatus with an input shaftconnected to said engine and an output shaft connected to said inputshaft of said transmission system, detecting respective rotating speedsof said input and said output shafts of said driving force transmissionapparatus, and detecting the transmission torque of said driving forcetransmission apparatus by using said detected rotating speeds asparameters.
 7. The discrimination method according to claim 6, includingproviding said driving force transmission apparatus with a hydro-dynamicpower transmitting device including pump means connected to said inputshaft of said driving force transmission apparatus and turbine meansconnected to said output shaft thereof, and arranging said hydro-dynamicpower transmitting device to transmit torque between said pump means andsaid turbine means by means of a working fluid.
 8. The discriminationmethod according to claim 5, including externally controlling saidtransmission torque of said driving force transmission apparatus, anddetecting said transmission torque by detecting the value of a controlparameter corresponding thereto.
 9. The discrimination method accordingto claim 8, including providing said driving force transmissionapparatus with clutch means disposed between said input and said outputshafts of said driving force transmission apparatus, and arranging theclutch means to transmit torque in accordance with a magnitude ofpressure of operating oil supplied thereto, and using as said controlparameter the pressure of said operating oil supplied to said clutchmeans.